Torsional vibration damper of a rotating shaft

ABSTRACT

This invention is directed at torsional vibration dampers of a rotating shaft. The dampers of the invention provide self-tuning to dampen harmonics, over a broad range of shaft angular velocities. Where the shaft rotates with an angular velocity about a longintudinal axis, and rotates perpendicular to a plane of rotation, the damper comprises: at least one passive damping element, and one controlling damping element.

FIELD OF THE INVENTION

The present invention relates generally to dampers for amelioratingtorsional vibrations of a rotating shaft, and more particularly toself-tuning dampers.

BACKGROUND OF THE INVENTION

Eliminating torsional vibration in a rotating shaft is an importantconsideration in the transmission of energy. Many modern systems rely ona rotating shaft to deliver kinetic energy from a motor (e.g. combustionengines). Such a shaft is susceptible to torsional vibrations having afrequency of vibration that is a natural multiple of the shaft angularvelocity. In many cases these vibrations arise in the motor, and are aresult of periodic combustion procedures therein. These vibrations areundesirable due to associated noise and/or equipment fatigue.

One known solution to the vibration problem is the use of a dual massflywheel. Such a flywheel acts as a pre-tuned resonant damper. However,a pre-tuned damper is not effective for dampening harmonics at variancewith its resonant frequency. This is not optimal, especially forundesired harmonics having frequencies that vary with shaft angularvelocity. Further, such a flywheel delays the response to the desiredgeneralized force of rotation.

U.S. Pat. No. 6,217,475 Dual-Mass Variable Inertia Flywheel Assemblydiscloses a vehicle driveline system includes a planetary geararrangement that cooperates with a flywheel member to provide a dualmass variable inertia flywheel assembly. The planetary gear arrangementpreferably includes an automated actuator that selectively engagesdifferent portions of the planetary gear arrangement to provide at leasttwo operating positions that each has an associated effective mass and acorresponding moment of inertia. Choosing an operating positionselectively controls the inertia of the flywheel assembly and greatlyreduces or eliminates torsion vibrations that may otherwise occur. Thefirst and second operating positions also provide high and low rangetransmission operation, respectively. This solution, however, does notprovide for a damper that is responsive to the varying shaft angularvelocity.

U.S. Pat. No. 6,192,851 Vibration Reducing System For InternalCombustion Engine discloses a vibration reducing system of an internalcombustion engine for an automotive vehicle. The vibration reducingsystem comprises a roll vibration system which generates a rollvibration of an engine main body, the roll vibration having a firstvibration mode. Additionally, a rotational vibration system is providedto generate a rotational vibration having a second vibration mode, andincludes a crankshaft of the engine, for generating a rotational drivingforce, a main flywheel fixedly connected to the crankshaft, a drivingforce transmitting mechanism through which the rotational driving forceof the crankshaft is transmitted, the driving force transmittingmechanism being movably secured to the engine main body, and an inertialmass member drivably connected to the driving force transmittingmechanism and rotatable to generate an inertial force upon receiving therotational driving force transmitted through the driving forcetransmitting mechanism. In the vibration reducing system, the first andsecond vibration modes cause anti-resonance at an anti-resonancefrequency. Additionally, the rotational vibration system is adjusted tocause the anti-resonance frequency to be generally coincident with oneof frequencies which are obtained respectively by multiplying anengine-revolution frequency at a predetermined engine speed by valueseach being represented by (a natural number/2). This solution isdirected at countering engine vibration, not torsion vibration of theshaft.

U.S. Pat. No. 5,733,218 Flywheel Having Two Centrifugal Masses And ATorsional Vibration Damper With Gear Train Elements Which Can BeAdjusted As A Function Of Load discloses a torsional vibration damperhas an input-side transmission element and an output-side transmissionelement, at least one of which is connected, preferably by means of aspring device, with at least one moment-transmitting element of a geartrain which acts between the two transmission elements. At least one ofthe moment-transmitting gear train elements can be aligned before thecreation of a connection to the corresponding element with some play, orclearance, at least in the radial direction, with respect to thiselement carrier, and after the application to the gear train of anadjustment moment which effects an alignment of this transmissionelement with respect to the other transmission elements as a function ofthe force by moving the two transmission elements into a predeterminedrelative position, the gear train element in question can be providedwith a connection to the element carrier which fixes the element in itsposition where it has been aligned as a function of the force. Thisdisclosure also fails to address dampers that are responsive to thevarying shaft angular velocity.

U.S. Pat. No. 5,720,248 (EP 586 973 B1) Torsional Tunable Coupling ForDiesel Engine Drive Shaft discloses an improved coupling assembly isprovided for transmitting rotary power from the working end of aninternal combustion engine to a driven shaft. The crankshaft also has afree end connected to an accessory drive train. The coupling assemblycomprises a low inertia flywheel having a mass so selected as to causethe node of the first crankshaft mode of torsional vibration to belocated in the vicinity of the middle of the crankshaft, and a flexiblecoupling which interconnects the working end of the crankshaft with thedriven shaft. The low inertia flywheel not only reduces the amplitude ofthe torsional deflection at the free end of the crankshaft, but furtherraises the primary torsional vibration orders of the engine which excitethe coupling assembly by at least one half of an order such that thepeak stresses applied to the teeth of the first gear wheel of theaccessory drive train are at least halved and are further applied to atleast twice as many gear teeth, thereby greatly prolonging the life ofthe gear wheel. Moreover, the low inertia flywheel further increases thelifetime of the flexible coupling by raising its natural frequency to alevel which is substantially higher than the 0.5 engine order oftorsional. vibration associated with engine malfunction and governorinteraction. This disclosure does not address dampers that areresponsive to the varying shaft angular velocity.

U.S. Pat. No. 5,570,615 Arrangement For Balancing Varying Moments AndVibrations In A Motor Vehicle Drive Train discloses technology tobalance varying moments and reduce vibrations in the drive train of amotor vehicle. Three flywheel masses and a clutch are provided in acommon housing. Two flywheel masses are connected by springs to providea dual-mass flywheel for reducing vibrations which can be transmittedinto the transmission from the crankshaft and the third flywheel mass isconnected to the first flywheel mass by planet gears mounted in fixedrelation to the housing so that the direction of rotation of the thirdflywheel mass is opposite to the direction of rotation of thecrankshaft. This disclosure is also directed at a pre-tuned solution,not at a responsive, self-tuning solution.

U.S. Pat. No. 5,493,936 Two-Mass Flywheel discloses a two-mass flywheelincluding a device that in series with a slip friction clutch couples asecondary mass rotationally-elastically to a primary mass. The slipfriction clutch includes two side disks, similar to cup springs, thatengage opposite faces of a center disk at the same effective diameter.The two side disks are preassembled to make a subassembly in which thespring preload can be adjusted after assembly. This is another pre-tunedsolution.

EP 349 624 Device For Suppressing Noise And Vibrations, In Particular InContinuously Variable Transmissions With Powers Split In Motor Vehiclesdiscloses a device for suppressing noise and vibration in continuouslyvariable transmissions with power split in motor vehicles. The power issplit into two branches, one of which passes through a continuouslyvariable converter. Said converter is insulated from the whole housingby means of special noise- and vibration-suppressing elements. Thenoise-producing torsional vibrations of the converter unit are alsoreduced to a minimum by a multi-weight flywheel system. This solutiondoes not propose to overcome the aforementioned limitations of apre-tuned multi-mass damper.

Application JP 2001057882 Vibration Damping Flywheel With Double Mass InAutomobile discloses a vibration damping flywheel with a double mass inan automobile capable of buffering shock of output torque of an engineand damping twisted vibration of a driving system by means of the doublemass. The vibration damping flywheel with the double mass in theautomobile comprises a flywheel with a primary mass and a secondarymass, a compressed coil spring, in which two sets of the compressed coilsprings respectively located in an inner side and outer side are filledup by grease, for improving damping effect in order to improve bucklingprevention and durable strength, a drive plate capable of flexiblyrealizing a hysteresis curve of the spring by compressing the coilspring while rolling on the compressed coil spring and by adjusting theshapes by means of an ax-shaped die or arch-shaped die, a roller forcompressing while the drive plate is rotating itself and also rolling onthe spring, and a spring guide capable of supporting a compression typecoil spring. This application does not address the issue of pre-tuningdeficiencies.

Applications JP 10018523 and JP 10018513 Device Having Buffer DeviceLaid Between At Least Two Flywheel Mass Bodies Capable Of RelativelyTurning For Each Other disclose making the turning resistance of abuffer device variable for the number of revolutions or a centrifugalforce, and installing a pressure accumulator effective in a peripheraldirection and a pressure accumulator effective in an axial direction.Pre-tuning limitations are also inherent in this technology.

Application JP10018523 discloses a flywheel wherein the axis of abscissarelates to a turning angle across the flywheel mass bodies, while theaxis of ordinate shows the transmissible moment of an elastic torsionalvibration damping device and a transmissible moment via a slide clutchrelates to the centrifugal force of a friction means. The coil spring ofthe damping device is somewhat compressed, due to the relative turningmotion of the flywheel mass bodies from the non-acting position thereof.Moment generated continues until becoming equal to the slide moment of afriction means. Then, this friction means slips to a turning angle, dueto turning in the same direction. When the means slips, exceeding theangle, the spring is further compressed, and a block is formed after thepassage of the means through a turning angle range. Furthermore, whenmoment exceeds a value transmissible via the friction means, theflywheel mass bodies become capable of giving a relative turning motion.As a result, the variation of the moment can be dampened or eliminated.This technology is not applicable to problem of per-tuning.

In application JP10018513, an intermediate plate of the output part of afriction slide clutch forms the input part of a buffer device, and thebuffer device has a first disc and a second disc at both sides of theintermediate plate. The first and second discs are connected to a rotarymass body at an axial gap in such a state as incapable of turning. Also,a pressure accumulation member made of a coil spring is housed in thewindows of the zone of the intermediate plate, and the coil spring actsagainst the relative rotation of the intermediate plate, and the firstand second discs. In addition, a friction device is laid between firstand second rotary mass bodies in parallel with the spring, and at aposition between the first disc and the zone of the first rotary massbody in an axial direction. Thereafter, the friction device is set andtightened between the disc and a crimp ring, and a friction ring is laidbetween the crimp ring and the zone in an axial direction. Also, thecrimp ring is peripherally fixed to the first disc. According to thisconstruction, a vibration can be dampened. This technology is alsoinapplicable to the problem at hand.

Application JP 06165698 Crankshaft Device For Two-Cylinder Four-CycleEngine, is directed to reduce energy loss even in the case that massiveerror occurs at balance weights when a crankshaft alone is balanced, byeasily adjusting the massive error afterward, and reducing the mass of asecond balance weight. A fan driving pulley and a flywheel are fixed toa crankshaft in which a crank angle of 180 degrees are shown between afirst crank pin and a second crank pin. A first balance weight whichsuppresses a vibration moment of the crank shaft is arranged on a crankarm on one end side of the crankshaft. A second balance weight isarranged on a flywheel fixed to the other end side of the crankshaft.This flywheel technology is not addressed to overcome pre-tuninglimitations.

Patent JP 2588838 Belt-Type Continuously Variable Transmission isdirected to improve vibration reducing performance of a flywheel anddurability of its supporting part by variably forming inertia mass ofthe flywheel in a belt type continuously variable transmission. A drivepulley, provided in a drive shaft connected to an engine through aflywheel, can be connected to the drive shaft through a forward clutch.When a shift range is placed in a stop range with small inertia massonly required for the flywheel, the forward clutch is disengaged. Whenthe shift range is placed in a running range with large inertia massrequired for the flywheel, the forward clutch is engaged, and inertiamass in a full line part in the drawing is used to serve as thesubstantial inertia mass of the flywheel. This technology is also notapplicable to overcome pre-tuning limitations.

U.S. Pat. No. 6,427,656, Internal Combustion Engine Including A Means OfReducing Cyclic Disturbances For Low-Speed Running discloses aninvention related to an internal combustion engine, the crankshaft ofwhich is equipped with a pulley or flywheel secured to it by fasteningmeans, in which the flywheel is equipped with at least one pendularelement, whose size, mass and position on the flywheel are determined soas to be tuned to close to the angular frequency of the major harmonicof the cyclic disturbance. This technology does address a self-tuningsolution. However, the pendulum solutions incorporated therein(Flyweights, Liquid mixtures, etc.) do not comprehensively address allacceleration patterns. Also, the disclosed unattached flyweights orliquid mixtures may not be compatible with other design considerationse.g. engine noise, environmental concerns, etc. Further, the solutiondisclosed addresses self-tuning multi-mass flywheels for single harmonicannihilation only.

Japanese application 11323776, Rolling Vibration Reducer For InternalCombustion Engine, discloses a rolling vibration reducer for an internalcombustion engine. The reducer offers a rolling vibration reducingeffect in specified operating conditions even if the inertial mass of amain flywheel system is changed. A rotor portion of a compressor for anair-conditioner, constituting a main flywheel system, is connected to acompressor pulley rotationally driven by an accessory driving belt viaan electromagnetic clutch in an approachable/separable manner. Arotating shaft of the rotor portion is connected to the electromagneticclutch and provided with a torsional spring portion of which springconstant is operated along the direction of the torsion of the rotatingshaft. In this way, in rotationally driving the rotor portion, with theoperation of the spring constant of the torsional spring portion, theantiresonant frequencies of the vibration of the main flywheel systemand the vibration of a sub-flywheel are kept to approximately agree withany of the frequency natural number/2 times of a rotational frequency atthe preset rotating speed of a crank shaft. Although this applicationdemonstrates the incorporation of an electromagnetic clutch, it onlyaddresses preset shaft speeds.

Thus, it is not known in the art to provide a damper for shaft torsionalvibrations capable of dampening multiple frequencies that are a functionof the shaft angular velocity. Nor is it known to similarly dampensingle frequencies, except by pendular free flyweights and liquidmixtures. Ideally, the response delay, to the desired generalized forceof rotation, of an improved damper is less than that of a dual massflywheel.

SUMMARY OF THE INVENTION

The prior art catalogued above fails to provide for multi-harmonicself-tuning. It also fails to provide an alternative to free flyweightsor liquid mixtures for single-harmonic self-tuning. Neither does theprior art address electromagetic solutions, except for preset shaftvelocities.

It is, thus, an object of an aspect of the invention to provide newself-tuning dampers for shafts rotatable over a broad variable range ofangular velocities.

Further it is an object of an aspect of the invention to provideself-tuning dampers for shafts rotatable over a broad variable range ofangular velocities, wherein the dampers respond to multiple harmonicfrequencies of the angular velocity.

According to one aspect of the present invention there is provided adamper for mitigating torsional vibrations of a shaft, rotating with anangular velocity about a longitudinal axis, and rotating perpendicularto a plane of rotation, said damper comprising:

at least one passive damping element, and

one controlling damping element.

According to another aspect of the present invention there is provided adamper for ameliorating torsional vibrations of a shaft, rotating withan angular velocity, said damper comprising:

a spring having a spring constant of proportionality with respect tomotion in a first degree of freedom,

a mass coupled to said spring for oscillation along said first degree offreedom, wherein said oscillation dampens said torsional vibrations ofsaid shaft corresponding to a frequency of said oscillation,

a selector coupled to said spring for movement along said spring in asecond degree of freedom, said movement under governance of said angularvelocity of said shaft, such that said spring constant of said spring,for said oscillation, is selected.

According to yet another aspect of the present invention there isprovided a damper for reducing torsional vibrations of a shaft, rotatingabout a longitudinal axis, and perpendicular to a plane of rotation,said damper comprising:

a joint, displaced from said longitudinal axis, and in said plane,

a pendulum having a degree of freedom for pendulum motion, about saidjoint, said motion in said plane, such that said pendulum motion dampenssaid torsional vibrations corresponding to the frequency of saidpendulum motion.

According to one aspect of the present invention there is provided adamper for reducing torsional vibrations of a rotating shaft, saiddamper comprising:

a first spring,

a second spring,

a mass physically coupled to said first spring and electromagneticallycoupled to said second spring for oscillation of a frequency, whereinsaid oscillation dampens said torsional vibrations of said shaft thatcorrespond to said frequency,

accelerometers coupled to the mass and the shaft for detecting therelative motion of said mass and said shaft,

a current generator for adjusting an electromagnetic bond whereby thesecond spring is coupled to the mass,

a computer coupled to said accelerometers and said current generator fordetecting at least one undesired said torsional vibration, determining acorresponding dampening spring stiffness improvement, and signalingcurrent generator to adjust current in order to implement saidimprovement.

According to still another aspect of the present invention there isprovided a method for damping torsional vibrations of a rotating shaftwherein said shaft includes a hub, a mass physically coupled to said hubvia a first spring and coupled to said hub via a second spring andelectromagnetic bond, said method comprising:

(i) Oscillating said mass angularly with respect to said hub in a mannerthat absorbs energy with a resonance related to the total effectivespring constants of the first and second springs,

(ii) identifying undesired harmonic motion, in said mass relative tosaid hub,

(iii) calculating applied current changes that, when applied by acurrent generator to said electromagnetic bond, change the totaleffective spring constant, and improve dampening of the detectedundesired harmonic motion, and

(iv) Applying said current changes.

According to still yet another aspect of the present invention there isprovided a damper for mitigating torsional vibrations of a shaft,rotating with an angular velocity, said damper comprising:

a spring having a spring constant of proportionality with respect tomotion in a first degree of freedom,

a mass coupled to said spring for (i) oscillation along said firstdegree of freedom, wherein said oscillation dampens said torsionalvibrations of said shaft corresponding to a frequency of saidoscillation, and (ii) movement along said spring in a second degree offreedom, said movement under governance of said angular velocity of saidshaft, such that said spring constant of said spring, for saidoscillation, is selected.

BRIEF DESCRIPTION OF THE DRAWINGS

Preferred embodiments of the present invention will now be described, byway of example only, with reference to the attached Figures, wherein:

FIG. 1 shows a perspective representation of a rotating shaft upon whicha damper according to the present invention may be implemented;

FIG. 2 shows an end view of the shaft of FIG. 1 with one embodiment ofthe invention;

FIG. 3 shows a partial end view of the shaft of FIG. 1 with springs andmasses according to an alternative embodiment of the invention;

FIG. 4 shows a top cross section of a spring from FIG. 3 providingadditional detail;

FIG. 5 shows an end view of the shaft of FIG. 1 with another embodimentof the invention in part (not all masses and springs shown);

FIG. 5A shows the end view of FIG. 5 with all masses and springs shown;

FIG. 6 shows an end view of the shaft of FIG. 1 with yet anotherembodiment of the invention;

FIG. 7 shows an end view of the shaft of FIG. 1 with still anotherembodiment of the invention;

FIG. 8 shows a schematic diagram of one electromagnetic apparatus forco-operating with the embodiment of FIG. 7; and

FIG. 9 shows a schematic diagram of an alternative electromagneticapparatus for co-operating with the embodiment of FIG. 7.

DETAILED DESCRIPTION OF THE INVENTION

Referring to FIG. 1, a shaft 105 rotates, with angular movementdisplacement θ and angular velocity $\begin{matrix}{{v = {\frac{\partial}{\partial t}\theta}},} & \quad\end{matrix}$about its primary axis 110, forced by the general force f_(θ) of momentabout axis 110, according to the prior art.

In conventional applications, the shaft 105 is incorporated in a motor,generator, engine, engine transmission, etc. In such applications, theshaft 105 is subjected to harmonic irregularities in the otherwisesmooth delivery of force f_(θ) to the shaft. Thus, the generalized forcef_(θ) of moment about axis 110 is a composite of desired force f_(θD)and irregular force f_(θI), comprising one or more undesired harmonicgeneralized forces, or $\begin{matrix}{{f_{\theta} = {f_{\theta D} + f_{\theta\quad l}}}{where}} & (1) \\{f_{\theta\quad l} = {\sum\limits_{n = 1}^{\infty}\quad\Gamma_{n}}} & (2)\end{matrix}$and Γ_(n) is a periodic function having a period T related to angularvelocity byv=2πn/T   (3)These irregular force harmonics result from the device (e.g. combustionengine) applying the force f_(θ) as it undergoes periodical procedures(combustion phases, valve operation, etc.). Of particular concern arethe first modes of irregular force harmonics (e.g. n=1, 2, 3, 4, 5, 6,7, 8), which are, typically, the only harmonics large enough to causeconcern.

In FIG. 2, there is shown, a damper 200, according to a first embodimentof the invention. The damper 200 addresses one of the undesired harmonicgeneralized forces f_(θI). Hub 205 corresponds to the shaft 105 of FIG.1, and is therefore shown integral herewith, but may be mounted thereon.The hub is connected to ring 214 via cantilever springs 212. Cantileversprings 216 are, likewise to springs 212, connected to ring 214, but arefree at the near hub end. Masses 218 are mounted freely on springs 216and coupled to the hub 205 by springs 220. Alternatively (not shown),springs 220 may couple masses 218 to ring 214 only, or to both ring 214and hub 205. Masses 218 are free to move along springs 216 (representedas r₂₁₈), except for the counter-forces due to the compression orextension of helical springs 220. Further, masses 218 are free to moveperpendicular to the radius (represented as θ₂₁₈), except for thecounter-forces due to displacement of springs 216. The entire apparatus200 is symmetrical, in order to ensure the moment of inertia is centeredabout hub 205. The maximum radial displacement of masses 218 isr_(218-max); The displacement of masses 218 in equilibrium position(v=0, springs 220 undistorted) is r₂₁₈₋₀.

Where m₂₁₈ is the mass of mass 218, and k₂₂₀ is the spring constant ofspring 220, the deviation of mass 218 from equilibrium position isr ₂₁₈ =r ₂₁₈₋₀ m ₂₁₈ v ² /K ₂₂₀   (4)

As the radial position of the mass 218 is extended, the length ofcantilever spring 216 for providing bending beam resistance iseffectively shortened. The shortening of the beam, in turn, implies agreater perpendicular force, for a perpendicular displacement i.e. asthe length of the bending beam is shortened, the spring constant ofproportionality relating displacement to force k₂₁₆ increases. Theseeffects, the outward displacement of masses 218 under angular velocityand the corresponding increase of the spring constant of proportionalityk₂₁₆ as the masses 218 move outward, combined with select dimensioningof the spring 216, result in a proportional relationship ofK ₂₁₆=3EJ/{L ³[1-(3r ₂₁₈/2L)+(r ₂₁₈ ³/2L ^(3)]}) K ₂₁₆ /m ₂₁₈ =v ²   (5)over the operating range of v. In these equations, E represents Young'sModulus, J represents the cross-section effectiveness of the cantileverand L represents the initial length of cantilever spring 216. In otherwords, masses 218 act as selectors for choosing k₂₁₆ in accordance withv. Thus, the equilibrium radial motion of masses 218 will have aconsequence regarding the angular motion of masses 218. Thus the radialmotion of mass 218 has the effect of governing factors affecting theangular simple harmonic oscillator motion of mass 218. The masses 218,in oscillating along θ₂₁₈, function as a tuned mass damping flywheel,where the spring characteristic k₂₁₆ varies according to (nv)². Thiseffectively dampens the torsion due to one irregular force harmonic.

The essence of this embodiment is:

(i) a spring (spring 216),

(ii) a mass (mass 218) coupled to the spring and oscillating in a firstdegree of freedom (θ₂₁₈), and

(iii) a selector (also, in this embodiment, mass 218) moving in a seconddegree of freedom (r₂₁₈) along the spring, under the governance of theshaft angular velocity v and, moving in a manner that determines theforce-displacement proportionality constant of the spring mentioned in(i) as that spring relates to the oscillating mass in (ii).

In other terms: As angular velocity v increases and the position of mass218 relocates to periphery, and there is a corresponding increase in themoment of inertia of the spring 216—mass 218 system coupled togetherwith re-adjusting and re-tuning in accordance to equilibrium (5). Thisphenomenon contributes to better damping characteristics of system.

Referring to FIG. 3, one spring 316, of another embodiment 300, isshown, similar to spring 216 of FIG. 2, for addressing three of theundesired harmonic generalized forces f_(θI). In this arrangement 300,the hub 305, connecting ring 314, springs 320, and 316, and mass 318 areas discussed with reference to FIG. 2, having regard to the hub 205,connecting ring 214, springs 220, and 216, and mass 218, respectively.

One arrangement 300 is shown in FIG. 3 for simplicity, although anotheridentical arrangement 300 is necessary on the opposite side of hub 305as per the symmetry shown in FIG. 2. In addition to features in commonwith the first embodiment 200, an additional helical spring 322 connectsmass 318 to mass 324. Mass 324 is mounted in a channel (slot) 332 ofspring 316, and includes three distinct components: a guide-joint 326,connected to a pendulum mass 330 by a stretching springing rod 328.Spring 316 also incorporates a pass through gap 334 (transverse to FIG.3) so that pendulum mass 330 is free to rotate about the guide-joint 326without striking the spring 316. Masses 318 and 324 are free to move inthe same manner as masses 218 of FIG. 2 i.e. they are capable of aradial and angular motion (r₃₁₈,θ₃₁₈) (r₃₂₄,θ₃₂₄). Mass 318 is governedby inertial rotational force proportional to v², given proper selectioncriteria as noted above. As with masses 218, mass 324 is influenced bythe angular velocity of the shaft v, and the force due to helical spring322. Provided the dimensions of the spring 316 are properly selected,the effective spring constant of spring 316 at the location (B) of mass324 is proportional in the same manner as for the case of spring 216 inthe previous embodiment. Assuming dimensions are chosen such that thesame harmonic is not compensated, mass 318 and 324, and theirsymmetrical counterparts dampen two of the irregular force harmonics.

A third irregular force harmonic is nullified by the pendulumoscillation of mass 324. Mass 324 undergoes pendulum motion about joint326, which is displaced from the center of hub 305. Under rotation ofthe hub 305 by v, mass 324, viewed from the rotating frame of reference,is subject to an apparent centrifugal forceF _(pendulum) =m ₃₂₄(v)² r ₃₂₄   (6)For oscillation of a pendulum through a small displacement under aconstant force (a reasonable approximation given the small oscillationdisplacement relative to the radius corresponding to angular motion θ)the pendulum's own period of oscillations isT ₃₂₆=2p(I ₃₂₈ /g)^(0.5)   (7)(approximating the location of the pendulum center of mass).), where gis acceleration of free fall (g=9.81 m/s²), and I₃₂₈ is (assuming thatr₃₂₈ is elongating, stretching of flexible rod 328 under additionalinertia force created by irregularities)I ₃₂₈ =k _(r)(I₃₂₈₋₀ +r ₃₂₈)   (8)

Referring to FIG. 4 there is shown a detailed side view 400 of thependulum formed of mass 324. In order for the flexible rod 328 andpendulum mass 330 to oscillate freely, a pass through cut 334 exists inspring 316. Flexible rod 328 includes means for flexing. A stiff helicalspring 329 longitudinally bisecting rod 328 is shown providing thisflexibility. However, other flexible means can be envisioned by oneskilled in the art. Guide-joint 326 of FIG. 3 includes guides 336, forcomplementing channels 338, in order that the guide remain properlyoriented in slot 332. Further, the guide-joint 326 has connectors 340that are connected to symmetrical springs 322 a and 322 b. These springsin combination serve the function of spring 322 of FIG. 3.

Referring to FIGS. 5 and 5A there is an end view of another embodimentof a damper 500 for dampening one undesired harmonic force. Hub 505 andouter ring 550 are co-centered, and rigidly fixed in relation to eachother. The superstructure (not shown) that makes this connection may besealed in order to ensure that the interior is protected from debris.Ring 514 is connected to the outer ring 550 by means of cantileversprings 552 coupled to ring 514 at cut through holes 554 in ring 514.These springs 552 are affixed to the outer ring 550, and generallyextend toward the hub 505. In order that ring 514 remain centered,struts 555 couple ring 514 to hub 505. Struts 555 extend outward fromhub 505. However, because ring 514 is free to oscillate under thegovernance of the cantilever springs 552, struts 555 include rollers 556at the coupling of ring 514 or, alternatively, an outer hub 558 coupledto hub 505 by roller bearings (not shown). In order to tune theoscillation of ring 514, tuning masses 562 slide along each of thecantilevers 552. The masses 562 are governed by helical springs 560,coupling adjacent masses 562 around the hub 505. Alternatively (notshown) springs may connect masses 562 in a radial direction to the outerhub 550 and/or the ring 514. Any of these arrangements generate the sametype of tuning as spring-masses 220 and 218 or 322 and 324 i.e. movementof a selector (masses 562) along a first degree of freedom, alters aspring constant (or constants) of proportionality resulting in a changein the resonant frequency of a simple harmonic oscillator along a seconddegree of freedom. In this damper 500, unlike the previous arrangements,the mass 514 moving in the second, tuned, degree of freedom, isdifferent than the selector (masses 562) moving in the first degree offreedom. However, either choice is within the scope of invention.

Referring to FIG. 6, there is shown an end view of yet anotherembodiment of a damper 600 for dampening one undesired harmonic force.

Elements of this embodiment 605, 614, 650, 652, 654, 655, 656, 658,correspond to elements 505, 514, 550, 552, 554, 555, 556, 558, of damper500 discussed above. Damper 600 differs from damper 500, however, in thesprings 682, 684 and mass 686 shown on one cantilever spring 652. Thesethree elements form an additional contour operating as per thespring-mass 218-220 combination of embodiment 200. The use of twosprings of this oscillating contour allows a greater selection and rangeof stiffness of the overall spring effect. Alternatively, mass 686 maybe disposed without a cantilever spring and one or both of the springs682, 684 may be two stiffness springs, having different stiffnesses withrespect to radial and angular motion.

Referring to FIG. 7, an arrangement 700 is shown for a self-tuning dualmass wherein the self-tuning is the result of electromagnetic feedback.As with the hub 505 and outer ring 550 of FIG. 5, there is a hub 705rigidly affixed to an outer ring 750. An inner ring 714 is co-centeredwith the hub 705 and connected by radial cantilever springs 712. Theouter ring is connected to cantilever springs 752. These springs 752 aredirected radially inward toward the inner ring 714 and terminate insheaths 770 that surround an arc of the ring 714 (like two inter-linkedannuli). Where the sheaths 770 and ring 714 have an applied current, anelectromagnetic bond is established between them. Where the currents arevaried, the degree of coupling is adjusted, thus providing for variancein spring forces (springs 712 plus some proportion of springs 752) thateffect oscillation of ring 712 about the hub 704.

FIG. 8 shows the electromagnetic apparatus 800 for providing feedback toarrangement 700. Accelerometers 884 are connected to each of the innerring 714 and the sheath 770 (shown in cross section). The accelerometers884 are each connected to integrators 890. Both outputs of Integrator890 are connected to computing system 896. The accelerometers 884 arealso connected to phase detectors 888 that are, in turn, connected toone phase discriminator 894. This is, in turn, connected to thecomputing system 896. Also connected to the sheath is a frequencydetector 882 that is connected, via filter 892, to the integrators 890and the computing system 896. Amplifiers and filters may be used, whereappropriate to ensure proper signal transmission. The computing systemis connected to a current generator 898, which is connected to the innerring 714 and sheath 770. The computing system 896 has the capacity to(i) input frequency, amplitude and phase difference data, (ii) performalgorithms on this data, and (iii) generate command signals for currentgenerators.

The computing system 896 accommodates the following algorithms, eitherof which provides for self-tuning:

In a first algorithm, (i) when the sheath oscillation amplitude isgreater in one interval, than the previous interval, and the inner ringamplitude is less than the amplitude of the previous interval (phase isnot equal to 90 degrees), the frequency must be reduced; (ii) when thesheath oscillation amplitude is less than, in one interval, theamplitude in the previous interval, and the inner ring amplitude isgreater than the amplitude of the previous interval (phase is not equalto 90 degrees), the frequency must be increased; (iii) when theamplitudes are constant from one interval to the next (phase equals 90degrees), the frequency remains constant.

Frequency change is accomplished by changing the current. When frequencyf is incremented by Δf, current j is incremented byΔj∝√{square root over ((Δf(2f+Δf))}   (9)and when frequency f is decremented by Δf, current j is decremented byΔj∝√{square root over ((Δf(2f−Δf))}.   (10)

In a second algorithm, (i) when the sheath oscillation amplitude in oneinterval is not equal to amplitude in previous interval the inner ringamplitude must be incremented simultaneously with (i)(a) a reduction offrequency if the sheath oscillation amplitude is greater in oneinterval, than the previous interval; or (i)(b) an increase of frequencyif the sheath oscillation amplitude is less in one interval, than theprevious interval; (ii) when the sheath amplitude is constant from oneinterval to the next, the frequency remains constant. Frequency changesare implemented as in the first algorithm. Subsequent to frequencyadjustment the inner ring 714 current amplitude is incremented unlessthe phase is equal to 90 degrees.

In a third algorithm, when the sheath oscillation amplitude is not equalto zero in one interval, an additional harmonic force is generated withfrequency equal to frequency of actual external force, and phase equalto −90° (−π/2). The amplitude of this force leads the sheath oscillationamplitude to a zero value.

Referring to FIG. 9, there is shown an alternative electromagneticapparatus 900 for providing feedback to arrangement 700. Accelerometers984 are connected to each of the inner ring 714 and the sheath 770(shown in cross section). The accelerometers 984 are each connected tospectrum analyzers 999 (e.g. Fourier analyzers), which are, in turn,connected to computing system 997. The computing system is connected toa current generator 998, which is connected to the inner ring 714 andsheath 770. Amplifiers and filters may be used, where appropriate, toensure proper signal transmission. The computing system 997 has thecapacity to (i) rank and filter input spectra to select undesiredharmonics, (ii) generate corresponding frequency and amplitude signals,(iii) determine phase difference between sheath and inner ringharmonics, (iv) perform algorithms on phase, frequency and amplitudedata, and (v) generate multi channel command signals for currentgenerators.

The aforementioned three algorithms for computing system 896, are alsoapplicable to step (iv) for computing system 997, with one modification.Apparatus 900 has the capacity to analyze and respond to multipleundesired harmonic frequencies, and therefore the algorithms can bemodified to operate for each of these harmonics independently. Thisresults in multiple channels for current generation.

For both apparatus and all algorithms, however, the operating principleis the same: Firstly, the mass of inner ring 714 is oscillating undersprings 712 and springs 752, where spring 752 is coupled to ring 714 byelectromagnetic forces due to applied current(s). This oscillationdampens torsional vibrations of shaft 705. Secondly, motion-to-signaltransducers (e.g. accelerometers) identify undesired harmonic motion, ininner ring 714 relative to sheath 770. Thirdly, calculations are made ontransducer output in order to determine an output that will yield acorresponding dampening spring stiffness improvement i.e. those appliedcurrent changes that, as a result of a change of total effective springconstant (for oscillation), improve dampening of the detected undesiredharmonic motion. Finally, those current changes are applied (feedback).The electromagnetic feedback solutions described herein amount tocontrolled passive, semi-active and active rather than passive elementsof the overall mechanical system, whereas the spring-mass systemsdescribed previously are passive elements.

Note that the various dampers disclosed herein are not mutuallyexclusive, and various permutations can be envisioned by one withordinary skill in the art, to address situations where a shaft has manyundesired torsional vibrations. A comprehensive example of such apermutation is the combination of a number of tunedspring-mass-cantilevers of FIG. 2, with a number of pendulums arrangedas per FIG. 3, and including a number of electromagnetically coupledrings as per FIG. 7 with integral spectral analysis. Using spectralanalysis alone there is no theoretical limit to the number of harmonicsthat can be addressed. However, it is believed to be more economical toaddress major harmonics with passive dampening as per FIGS. 2 or 3.

While the present invention has been described in detail for purposes ofimplementation, the above-described embodiments of the invention aremerely examples. Alterations and modifications may be effected thereto,by those of skill in the art, without departing from the scope of theinvention which is defined solely by the claims appended hereto.

1. (canceled)
 2. A damper for mitigating torsional vibrations of ashaft, rotating with an angular velocity about a longitundinal axis, androtating perpendicular to a plane of rotation, said damper comprising:at least one passive damping element, and at least one controllingdamping element, wherein said at least one passive damping elementcomprises: a spring having a spring constant of proportionality withrespect to motion in a first degree of freedom, a mass coupled to saidspring for oscillation along said first degree of freedom, wherein saidoscillation dampens said torsional vibrations of said shaftcorresponding to a frequency of said oscillation, a selector coupled tosaid spring for movement along said spring in a second degree offreedom, said movement under governance of said angular velocity of saidshaft, such that said spring constant of said spring, for saidoscillation, is selected.
 3. A damper for mitigating torsionalvibrations of a shaft, rotating with an angular velocity about alongitundinal axis, and rotating perpendicular to a plane of rotation,said damper comprising: at least one passive damping element, and atleast one controlling damping element, wherein said at least one passivedamping element comprises: a joint, displaced from said longitudinalaxis, and in said plane, a pendulum having a degree of freedom forpendulum motion, about said joint, said motion in said plane, such thatsaid pendulum motion dampens said torsional vibrations corresponding tothe frequency of said pendulum motion.
 4. A damper for amelioratingtorsional vibrations of a shaft, rotating with an angular velocity, saiddamper comprising: a spring having a spring constant of proportionalitywith respect to motion in a first degree of freedom, a mass coupled tosaid spring for oscillation along said first degree of freedom, whereinsaid oscillation-dampens said torsional vibrations of said shaftcorresponding to a frequency of said oscillation, a selector coupled tosaid spring for movement along said spring in a second degree offreedom, said movement under governance of said angular velocity of saidshaft, such that said spring constant of said spring, for saidoscillation, is selected.
 5. The damper of claim 4 wherein saidfrequency is a constant multiple of said angular velocity.
 6. The damperof claim 5 wherein said frequency is a constant multiple of said angularvelocity by a natural number.
 7. (canceled)
 8. A damper for reducingtorsional vibrations of a shaft, rotating about a longitundinal axis,and perpendicular to a plane of rotation, said damper comprising: ajoint, displaced from said longitudinal axis, and in said plane, apendulum having a degree of freedom for pendulum motion, about saidjoint, said motion in said plane, such that said pendulum motion dampenssaid torsional vibrations corresponding to the frequency of saidpendulum motion.
 9. The damper of claim 8 wherein said frequency is aconstant multiple of the angular velocity of said shaft.
 10. The damperof claim 9 wherein said frequency is a constant multiple of the angularvelocity of said shaft by a natural number.
 11. (canceled) 12.(canceled)
 13. (canceled)
 14. (canceled)
 15. (canceled)
 16. (canceled)17. (canceled)
 18. (canceled)
 19. (canceled)
 20. (canceled) 21.(canceled)
 22. A damper for mitigating torsional vibrations of a shaft,rotating with an angular velocity about a longitundinal axis, androtating perpendicular to a plane of rotation, said damper comprising:at least one passive damping element, and at least one controllingdamping element, wherein said at least one passive damping elementcomprises: wherein said at least one passive damping element comprises:a spring having a spring constant of proportionality with respect tomotion in a first degree of freedom, a mass coupled to said spring for(i) oscillation along said first degree of freedom, wherein saidoscillation dampens said torsional vibrations of said shaftcorresponding to a frequency of said oscillation, and (ii) movementalong said spring in a second degree of freedom, said movement undergovernance of said angular velocity of said shaft, such that said springconstant of said spring, for said oscillation, is selected.
 23. A damperfor mitigating torsional vibrations of a shaft, rotating with an angularvelocity, said damper comprising: a spring having a spring constant ofproportionality with respect to motion in a first degree of freedom, amass coupled to said spring for (i) oscillation along said first degreeof freedom, wherein said oscillation dampens said torsional vibrationsof said shaft corresponding to a frequency of said oscillation, and (ii)movement along said spring in a second degree of freedom, said movementunder governance of said angular velocity of said shaft, such that saidspring constant of said spring, for said oscillation, is selected. 24.The damper of claim 22 wherein said frequency is a constant multiple ofsaid angular velocity.
 25. The damper of claim 23 wherein said frequencyis a constant multiple of said angular velocity by a natural number. 26.A damper for mitigating torsional vibrations of a shaft, rotating withan angular velocity about a longitundinal axis, and rotatingperpendicular to a plane of rotation, said damper comprising: at leastone passive damping element, and at least one controlling dampingelement, wherein said at least one passive damping element comprises:said at least one passive damping element comprises: at least twosprings having two spring constants of proportionality with respect tomotion in a first degree of freedom, a mass coupled to said two springsfor oscillation along said first degree of freedom, wherein saidoscillation dampens said torsional vibrations of said shaftcorresponding to a frequency of said oscillation, and wherein said massis coupled to said two springs for movement in a second degree offreedom, said movement under governance of said angular velocity of saidshaft, such that the moment of inertia of said mass about said shaft isgoverned by said angular velocity.
 27. A damper for mitigating torsionalvibrations of a shaft, rotating with an angular velocity about alongitundinal axis, and rotating perpendicular to a plane of rotation,said damper comprising: at least one passive damping element, and atleast one controlling damping element, wherein the controlling dampingelement is capable of damping at least one vibration of a frequencydifferent from any resonant frequencies of the passing damping element.28. The damper of claim 4, wherein said selector is said mass.
 29. Thedamper of claim 5, wherein said selector is said mass.
 30. The damper ofclaim 6, wherein said selector is said mass.